Note: Descriptions are shown in the official language in which they were submitted.
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TITLE OF THE INVENTION
COAXIAL ECONOMIZER ASSEMBLY AND METHOD
CROSS-REFERENCE TO RELATED APPLICATIONS
[0001] Not applicable presently.
FEDERALLY SPONSORED RESEARCH OR DEVELOPMENT
[0002] None.
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BACKGROUND OF THE INVENTION
[0003] The present invention generally relates to an economizer for flash
cooling a
refrigerant liquid, and specifically with an economizer arranged coaxially
with a condenser or
other structure, e.g. an evaporator, for use in a refrigeration system having
at least two stages
of compression.
[0004] Refrigeration systems typically incorporate a refrigeration loop
to provide
chilled water for cooling a designated building space. A typical refrigeration
loop includes a
compressor to compress refrigerant gas, a condenser to condense the compressed
refrigerant
to a liquid, and an evaporator that utilizes the liquid refrigerant to cool
water. The chilled
water is then piped to the space to be cooled.
[0005] One such refrigeration or air conditioning system uses at least
one centrifugal
compressor and is referred to as a centrifugal chiller. Centrifugal
compression involves the
purely rotational motion of only a few mechanical parts. A single centrifugal
compressor
chiller, sometimes called a simplex chiller, typically range in size from 100
to above 2,000
tons of refrigeration. Typically, the reliability of centrifugal chillers is
high, and the
maintenance requirements are low.
[0006] Centrifugal chillers consume significant energy resources in
commercial and
other high cooling and/or heating demand facilities. Such chillers can have
operating lives of
upwards of thirty years or more in some cases.
[0007] Centrifugal chillers provide certain advantages and efficiencies
when used in a
building, city district (e.g. multiple buildings) or college campus, for
example. Such chillers
are useful over a wide range of temperature applications including Middle East
conditions.
At lower refrigeration capacities, screw, scroll or reciprocating-type
compressors are most
often used in, for example, water-based chiller applications.
[0008] One component of existing chillers is an economizer. The
economizer
improves the operating efficiency of the system.
[0009] An economizer is typically utilized between the condenser and the
evaporator
of a refrigeration system to cool refrigerant liquid below the temperature at
which it leaves
the condenser. Flash cooling is achieved by the evaporation of part of the
refrigerant liquid
as it flows from the condenser through nozzles, orifices, or other pressure
reducing means
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into a chamber which is lower in pressure. The flashing refrigerant cools the
remaining liquid
by absorbing heat as it vaporizes. Upon separation from the cooled liquid, the
refrigerant
vapor, or flash gas, is conveyed to the inlet of a compressor stage operating
at intermediate
pressure. The cooled refrigerant liquid flows from the economizer to an
evaporator, where it
is vaporized in heat exchange relationship with another fluid, e.g., water, to
satisfy a cooling
load. Refrigerant vapor leaving the evaporator is typically compressed in two
or more stages
of compression. Prior economizers have been designed as separate units,
distinct from the
condenser, compressor and other structures common to chiller systems.
[0010] Prior chiller designs also typically connect the first stage
discharge of a
compressor to a second stage compressor and include complicated casting and
piping
arrangements. These designs are sometimes called two-stage, in-line designs.
[0011] Essentially, these in-line designs have a series of continuous
castings that
allow the discharge gas leaving a first stage compressor to be delivered into
the inlet of the
second stage compressor. The impeller of the first stage compressor imposes a
great deal of
tangential velocity to the fluid being compressed. This fluid with a
tangential velocity is
called swirling flow. As the fluid flows through the diffuser of the first
stage compressor, it
passes through a 180 U-bend. A set of blades in the return channel bend are
typically used
in an attempt to direct the fluid flow in an axial direction at the inlet to
the second stage
compressor. This swirling fluid flow is combined with the flash gas flow from
the
economizer to essentially inter-cool the swirling gas of the first stage
compression. In
practice, the mixing of the two flows is not as thorough as desired and
predominately occurs
far enough down the fluid flow path, e.g. in the impellers of the second
stage, that only a
modest efficiency improvement is gained.
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BRIEF SUMMARY OF THE INVENTION
[0012] According to a preferred embodiment of the present invention, a
coaxial
economizer for use in a chiller system comprises an inner housing and an outer
housing
having a common longitudinal axis. The outer housing has an inlet for
receiving a fluid
from an upstream compressor stage of a multistage compressor and an outlet for
conveying a fluid to a downstream compressor stage of a multistage compressor.
A flow
chamber forms a fluid flow path about the inner housing. The coaxial
economizer further
comprises a flash chamber for flashing fluid in a liquid state to a gas state.
A flow
passage between said flash chamber and the flow chamber conveys a flashed gas
from the
flash chamber to the flow chamber. The flashed gas conveyed from the flash
chamber
and the fluid received from the inlet of the outer housing mixes along the
fluid flow path
toward the outlet of the outer housing.
[0013] In yet another preferred embodiment of the present invention, a
method of
flowing fluid through a coaxial economizer in a chiller system comprises the
steps of:
receiving a fluid from an upstream compressor stage of a multistage compressor
into a
coaxial economizer; flashing a liquid to gas within a flash chamber of the
coaxial
economizer; passing the gas within the flash chamber through a flow passage to
the flow
chamber of the coaxial economizer; and mixing and flowing the gas conveyed
from the
flash chamber and the fluid received from the inlet of the outer housing along
the fluid
flow path to the outlet of the coaxial economizer. The coaxial economizer of
this method
comprises: an inner housing and an outer housing having a common longitudinal
axis;
said outer housing having an inlet for receiving a fluid from the upstream
compressor
stage and an outlet for conveying a fluid to a downstream compressor stage of
a
multistage compressor; a flow chamber forming a fluid flow path about the
inner
housing; a flash chamber for flashing fluid in a liquid state to a gas state;
and a flow
passage between said flash chamber and the flow chamber for conveying a
flashed gas
from the flash chamber to the flow chamber; wherein the flashed gas conveyed
from the
flash chamber and the fluid received from the inlet of the outer housing mix
along the
fluid flow path toward the outlet of the outer housing.
[0014] Embodiments of the coaxial economizer eliminates the traditional
in-line
design, combines multiple functions into one integrated system, improves fluid
mixing of
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the inter-cooled gas prior to entry of the second stage and improves fluid
flow dynamics
(e.g. swirl reduction) through the system, which, in turn, improves chiller
performance.
The
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coaxial economizer is operable over a wide capacity range, and provides
improved efficiency
in a compact size.
[0015] Additional advantages and features of the invention will become
more
apparent from the description of a preferred embodiment of the present
invention and the
claims which follow.
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BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWINGS
[0016] The following figures include like numerals indicating like
features where
possible:
[0017] Figure 1 illustrates a perspective view of a chiller system and
the various
components according to an embodiment of the present invention.
[0018] Figure 2 illustrates an end, cut away view of a chiller system
showing tubing
arrangements for the condenser and evaporator according to an embodiment of
the present
invention.
[0019] Figure 3 illustrates another perspective view of a chiller system
according to
an embodiment of the present invention.
[0020] Figure 4 illustrates a cross-sectional view of a multi-stage
centrifugal
compressor for a chiller system according to an embodiment of the present
invention.
[0021] Figure 5 illustrates a perspective view of a conformal draft pipe
attached to a
coaxial economizer arrangement according to an embodiment of the present
invention.
[0022] Figure 6 illustrates a view of a swirl reducer and vortex fence
positioned in a
first leg of a three leg suction pipe between a conformal draft pipe attached
to a coaxial
economizer arrangement upstream of a final stage compressor according to an
embodiment of
the present invention.
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DETAILED DESCRIPTION OF A PREFERRED EMBODIMENT
[0023] Referring to FIGS. 1-3 of the drawings, a chiller or chiller
system 20 for a
refrigeration system. A single centrifugal chiller system and the basic
components of chiller
20 are illustrated in FIGS. 1-3. The chiller 20 includes many other
conventional features not
depicted for simplicity of the drawings. In addition, as a preface to the
detailed description, it
should be noted that, as used in this specification and the appended claims,
the singular forms
"a," "an," and "the" include plural referents, unless the context clearly
dictates otherwise.
[0024] In the embodiment depicted, chiller 20 is comprised of an
evaporator 22,
multi-stage compressor 24 having a non-final stage compressor 26 and a final
stage
compressor 28 driven by a variable speed, direct drive permanent magnet motor
36, and a
coaxial economizer 40 with a condenser 44. The chiller 20 is directed to
relatively large
tonnage centrifugal chillers in the range of about 250 to 2000 tons or larger.
[0025] In a preferred embodiment, the compressor stage nomenclature
indicates that
there are multiple distinct stages of gas compression within the chiller's
compressor portion.
While a multi-stage compressor 24 is described below as a two-stage
configuration in a
preferred embodiment, persons of ordinary skill in this art will readily
understand that
embodiments and features of this invention are contemplated to include and
apply to, not
only two-stage compressors/chillers, but to single stage and other multiple
stage
compressors/chillers, whether in series or in parallel.
[0026] Referring to FIGS. 1-2, for example, preferred evaporator 22 is
shown as a
shell and tube type. Such evaporators can be of the flooded type. The
evaporator 22 may be
of other known types and can be arranged as a single evaporator or multiple
evaporators in
series or parallel, e.g. connecting a separate evaporator to each compressor.
As explained
further below, the evaporator 22 may also be arranged coaxially with an
economizer 42. The
evaporator 22 can be fabricated from carbon steel and/or other suitable
material, including
copper alloy heat transfer tubing.
[0027] A refrigerant in the evaporator 22 performs a cooling function. In
the
evaporator 22, a heat exchange process occurs, where liquid refrigerant
changes state by
evaporating into a vapor. This change of state, and any superheating of the
refrigerant vapor,
causes a cooling effect that cools liquid (typically water) passing through
the evaporator
tubing 48 in the evaporator 22. The evaporator tubing 48 contained in the
evaporator 22 can
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be of various diameters and thicknesses and comprised typically of copper
alloy. The tubes
may be replaceable, are mechanically expanded into tube sheets, and externally
finned
seamless tubing.
[0028] The chilled or heated water is pumped from the evaporator 22 to an
air
handling unit (not shown). Air from the space that is being temperature
conditioned is drawn
across coils in the air handling unit that contains, in the case of air
conditioning, chilled
water. The drawn-in air is cooled. The cool air is then forced through the air
conditioned
space, which cools the space.
[0029] Also, during the heat exchange process occurring in the evaporator
22, the
refrigerant vaporizes and is directed as a lower pressure (relative to the
stage discharge) gas
through a non-final stage suction inlet pipe 50 to the non-final stage
compressor 26. Non-
final stage suction inlet pipe 50 can be, for example, a continuous elbow or a
multi-piece
elbow.
[0030] A three-piece elbow is depicted in an embodiment of non-final
stage suction
inlet pipe 50 in FIGS. 1-3, for example. The inside diameter of the non-final
stage suction
inlet pipe 50 is sized such that it minimizes the risk of liquid refrigerant
droplets being drawn
into the non-final stage compressor 26. For example, the inside diameter of
the non-final
stage suction inlet pipe 50 can be sized based on, among things, a limit
velocity of 60 feet per
second for a target mass flow rate, the refrigerant temperature and a three-
piece elbow
configuration. In the case of the multi-piece non-final stage suction inlet
pipe 50, the lengths
of each pipe piece can also be sized for a shorter exit section to, for
example, minimize
corner vortex development.
[0031] To condition the fluid flow distribution delivered to the non-
final stage
compressor 26 from the non-final stage suction inlet pipe 50, a swirl reducer
or deswirler
146, as illustrated in FIG. 6 and described further below, can be optionally
incorporated into
the non-final stage suction inlet pipe 50. The refrigerant gas passes through
the non-final
stage suction inlet pipe 50 as it is drawn by the multi-stage centrifugal
compressor 24, and
specifically the non-final stage centrifugal compressor 26.
[0032] Generally, a multi-stage compressor compresses refrigerant gas or
other
vaporized fluid in stages by the rotation of one or more impellers during
operation of the
chiller's closed refrigeration circuit. This rotation accelerates the fluid
and in turn, increases
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the kinetic energy of the fluid. Thereby, the compressor raises the pressure
of fluid, such
as refrigerant, from an evaporating pressure to a condensing pressure. This
arrangement
provides an active means of absorbing heat from a lower temperature
environment and
rejecting that heat to a higher temperature environment.
[0033] Details of the structure, function and operation of a preferred
compressor
assembly, which may include a mixed flow impeller and/or an inlet flow
conditioning
assembly, are disclosed in U.S. patent nos. 8,037,713 and 7,856,834 and U.S.
patent
application publication no. 2009/0208331, commonly assigned to the assignee of
the
present invention. A brief discussion of a preferred compressor assembly
follows;
however, other compressor assemblies may be used with embodiments of the
present
invention.
[0034] Referring now to FIG. 4, the compressor 24 is typically an
electric motor
driven unit. A variable speed drive system drives the multi-stage compressor.
The
variable speed drive system comprises a permanent magnet motor 36 located
preferably
in between the non-final stage compressor 26 and the final stage compressor 28
and a
variable speed drive 38 having power electronics for low voltage (less than
about 600
volts), 50 Hz and 60 Hz applications. The variable speed drive system
efficiency, line
input to motor shaft output, preferably can achieve a minimum of about 95
percent over
the system operating range.
[0035] While conventional types of motors can be used with and benefit
from
embodiments of the present invention, a preferred motor is a permanent magnet
motor
36. Permanent magnet motor 36 can increase system efficiencies over other
motor types.
[0036] A preferred motor 36 comprises a direct drive, variable speed,
hermetic,
permanent magnet motor. The speed of the motor 36 can be controlled by varying
the
frequency of the electric power that is supplied to the motor 36. The
horsepower of
preferred motor 36 can vary in the range of about 125 to about 2500
horsepower.
[0037] The permanent magnet motor 36 is under the control of a variable
speed
drive 38. The permanent magnet motor 38 of an embodiment is compact,
efficient,
reliable, and relatively quieter than conventional motors. As the physical
size of the
compressor assembly is reduced, the compressor motor used must be scaled in
size to
fully realize the benefits of improved fluid flow paths and compressor element
shape and
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size. Motor 36 is reduced in volume by approximately 30 to 50 percent or more
when
compared to conventional existing
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designs for compressor assemblies that employ induction motors and have
refrigeration
capacities in excess of 250-tons. The resulting size reduction of embodiments
of the present
invention provides a greater opportunity for efficiency, reliability, and
quiet operation
through use of less material and smaller dimensions than can be achieved
through more
conventional practices.
[0038] Typically, an AC power source (not shown) will supply multiphase
voltage
and frequency to the variable speed drive 38. The AC voltage or line voltage
delivered to the
variable speed drive 38 will typically have nominal values of 200V, 230V,
380V, 415V,
480V, or 600V at a line frequency of 50Hz or 60Hz depending on the AC power
source.
[0039] The permanent magnet motor 36 comprises a rotor 68 and a stator
70. The
stator 70 consists of wire coils formed around laminated steel poles, which
convert variable
speed drive applied currents into a rotating magnetic field. The stator 70 is
mounted in a
fixed position in the compressor assembly and surrounds the rotor 68,
enveloping the rotor
with the rotating magnetic field. The rotor 68 is the rotating component of
the motor 36 and
consists of a steel structure with permanent magnets, which provide a magnetic
field that
interacts with the rotating stator magnetic field to produce rotor torque. The
rotor 68 may
have a plurality of magnets and may comprise magnets buried within the rotor
steel structure
or be mounted at the rotor steel structure surface. The rotor 68 surface mount
magnets are
secured with a low loss filament, metal retaining sleeve or by other means to
the rotor steel
support. The performance and size of the permanent magnet motor 36 is due in
part to the
use of high energy density permanent magnets.
[0040] Permanent magnets produced using high energy density magnetic
materials, at
least 20 MGOe (Mega Gauss Oersted), produce a strong, more intense magnetic
field than
conventional materials. With a rotor that has a stronger magnetic field,
greater torques can be
produced, and the resulting motor can produce a greater horsepower output per
unit volume
than a conventional motor, including induction motors. By way of comparison,
the torque
per unit volume of permanent magnet motor 36 is at least about 75 percent
higher than the
torque per unit volume of induction motors used in refrigeration chillers of
comparable
refrigeration capacity. The result is a smaller sized motor to meet the
required horsepower
for a specific compressor assembly.
[0041] Further manufacturing, performance, and operating advantages and
disadvantages can be realized with the number and placement of permanent
magnets in the
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rotor 68. For example, surface mounted magnets can be used to realize greater
motor
efficiencies due to the absence of magnetic losses in intervening material,
ease of
manufacture in the creation of precise magnetic fields, and effective use of
rotor fields to
produce responsive rotor torque. Likewise, buried magnets can be used to
realize a simpler
manufactured assembly and to control the starting and operating rotor torque
reactions to load
variations.
[0042] The bearings, such as rolling element bearings (REB) or
hydrodynamic
journal bearings, can be oil lubricated. Other types of bearings can be oil-
free systems. A
special class of bearing which is refrigerant lubricated is a foil bearing and
another uses REB
with ceramic balls. Each bearing type has advantages and disadvantages that
should be
apparent to those of skill in the art. Any bearing type that is suitable of
sustaining rotational
speeds in the range of about 2,000 to about 20,000 RPM may be employed.
[0043] The rotor 68 and stator 70 end turn losses for the permanent
magnet motor 36
are very low compared to some conventional motors, including induction motors.
The motor
36 therefore may be cooled by means of the system refrigerant. With liquid
refrigerant only
needing to contact the stator 70 outside diameter, the motor cooling feed
ring, typically used
in induction motor stators, can be eliminated. Alternatively, refrigerant may
be metered to
the outside surface of the stator 70 and to the end turns of the stator 70 to
provide cooling.
[0044] The variable speed drive 38 typically will comprise an electrical
power
converter comprising a line rectifier and line electrical current harmonic
reducer, power
circuits and control circuits (such circuits further comprising all
communication and control
logic, including electronic power switching circuits). The variable speed
drive 38 will
respond, for example, to signals received from a microprocessor (also not
shown) associated
with the chiller control panel 182 to increase or decrease the speed of the
motor by changing
the frequency of the current supplied to motor 36. Cooling of motor 36 and/or
the variable
speed drive 38, or portions thereof, may be by using a refrigerant circulated
within the chiller
system 20 or by other conventional cooling means. Utilizing motor 36 and
variable speed
drive 38, the non-final stage compressor 26 and a final stage compressor 28
typically have
efficient capacities in the range of about 250-tons to about 2,000-tons or
more, with a full
load speed range from approximately 2,000 to above about 20,000 RPM.
[0045] With continued reference to FIG. 4 and turning to the compressor
structure,
the structure and function of the non-final or upstream stage compressor 26,
final or
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downstream stage compressor 28 and any intermediate stage compressor (not
shown) are
substantially the same, if not identical, and therefore are designated
similarly as illustrated in
the FIG. 4, for example. Differences, however, between the compressor stages
exist in a
preferred embodiment and will be discussed below. Features and differences not
discussed
should be readily apparent to one of ordinary skill in the art.
[0046] Preferred non-final stage compressor 26 has a compressor housing
30 having
both a compressor inlet 32 and a compressor outlet 34. The non-final stage
compressor 26
further comprises an inlet flow conditioning assembly 54, a non-final stage
impeller 56, a
diffuser 112 and a non-final stage external volute 60.
[0047] The non-final stage compressor 26 can have one or more rotatable
impellers
56 for compressing a fluid, such as refrigerant. Such refrigerant can be in
liquid, gas or
multiple phases and may include R-123 refrigerant. Other refrigerants, such as
R-134a, R-
245fa, R-141b and others, and refrigerant mixtures are contemplated. Further,
the present
invention contemplates use of azeotropes, zeotropes and/or a mixture or blend
thereof that
have been and are being developed as alternatives to commonly used
contemplated
refrigerants.
[0048] By the use of motor 36 and variable speed drive 38, multistage
compressor 24
can be operated at lower speeds when the flow or head requirements on the
chiller system do
not require the operation of the compressor at maximum capacity, and operated
at higher
speeds when there is an increased demand for chiller capacity. That is, the
speed of motor 36
can be varied to match changing system requirements which results in
approximately 30
percent more efficient system operation compared to a compressor without a
variable speed
drive. By running compressor 24 at lower speeds when the load or head on the
chiller is not
high or at its maximum, sufficient refrigeration effect can be provided to
cool the reduced
heat load in a manner which saves energy, making the chiller more economical
from a cost-
to-run standpoint and making chiller operation extremely efficient as compared
to chillers
which are incapable of such load matching.
[0049] Referring still to FIGS. 1-4, refrigerant is drawn from the non-
final stage
suction piping 50 to an integrated inlet flow conditioning assembly 54 of the
non-final stage
compressor 26. The integrated inlet flow conditioning assembly 54 comprises an
inlet flow
conditioning housing 72 that forms a flow conditioning channel 74 with flow
conditioning
channel inlet 76 and flow conditioning channel outlet 78. The channel 74 is
defined, in part,
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by a shroud wall 80 having an inside shroud side surface 82, a flow
conditioning nose 84, a
strut 86, a flow conditioning body 92 and a plurality of inlet guide
blades/vanes 100. These
structures, which may be complimented with swirl reducer 146, cooperate to
produce fluid
flow characteristics that are delivered into the vanes 100, such that less
turning of the vanes
100 is required to create the target swirl distribution for efficient
operation in impellers 56,
58.
[0050] The drawing of FIG. 4 also depicts a double-ended shaft 66 that
has a non-
final stage impeller 56 mounted on one end of the shaft 66 and a final stage
impeller 58 on
the other end of the shaft 66. The double-ended shaft configuration of this
embodiment
allows for two or more stages of compression. The impeller shaft 66 is
typically dynamically
balanced for vibration reduced operation, preferably and predominantly
vibration free
operation.
[0051] Different arrangements and locations of the impellers 56, 58;
shaft 66 and
motor 36 should be apparent to one of ordinary skill in the art as being
within the scope of the
invention. It should be also understood that in this embodiment the structure
and function of
the impeller 56, impeller 58 and any other impellers added to the compressor
24 are
substantially the same, if not identical. However, impeller 56, impeller 58
and any other
impellers may have to provide different flow characteristics impeller to
impeller.
[0052] In a preferred embodiment, fluid is delivered from the impellers
56, 58 and
diffusers 112 to a non-final stage external volute 60 and a final stage
external volute 62,
respectively for each stage. The volutes 60, 62, illustrated in FIG. 1-4, are
external. The
volutes 60, 62 have a centroid radius that is greater than the centroid radius
at the exit of the
diffuser 112. Volutes 60, 62 have a curved funnel shape and increase in area
to a discharge
port 64 for each stage, respectively. Volutes that lie off the meridional
diffuser centerline are
sometimes called overhung.
[0053] The external volutes 60, 62 of this embodiment replace the
conventional return
channel design and are comprised of two portions ¨ the scroll portion and the
discharge conic
portion. Use of volutes 60, 62 lowers losses as compared to return channels at
part load and
have about the same or less losses at full load. As the area of the cross-
section increases, the
fluid in the scroll portion of the volutes 60, 62 is at about a constant
static pressure so it
results in a distortion free boundary condition at the diffuser exit. The
discharge conic
increases pressure when it exchanges kinetic energy through the area increase.
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[0054] In the case of the non-final stage compressor 26 of this
embodiment, fluid is
delivered from the external volute 60 to a coaxial economizer 40. In the case
of the final
stage compressor 28 of this embodiment, the fluid is delivered from the
external volute 62 to
a condenser 44 (which may be arranged coaxially with an economizer).
[0055] Turning now to the coaxial economizer 40, the coaxial economizer
40 has an
economizer 42 arranged coaxially with a condenser 44. Applicants refer to this
arrangement
as an exemplary coaxial economizer 40. The coaxial economizer 40 combines
multiple
functions into one integrated system and further increases system
efficiencies. Coaxial is
used in the common sense where one structure (e.g. economizer 42) has a
coincident axis
with at least one other structure (e.g. the condenser 44 or evaporator 22). A
discussion of a
preferred coaxial economizer 40 follows.
[0056] By the use of coaxial economizer 40, additional efficiencies are
added to the
compression process that takes place in chiller 20 and the overall efficiency
of chiller 20 is
increased. The coaxial economizer 40 combines multiple functions into one
integrated
system and further increases system efficiencies.
[0057] Other coaxial economizer arrangements within the scope of this
invention
should be apparent. For example, while economizer 42 surrounds and is coaxial
with
condenser 44 in a preferred embodiment, it will be understood by those skilled
in the art that
it may be advantageous in certain circumstances for economizer 42 to surround
evaporator
22. An example of such a circumstance is one in which, due to the particular
application or
use of chiller 20, it is desired that evaporator 22, when surrounded by
economizer 42, acts, in
effect, as a heat sink to provide additional interstage cooling to the
refrigerant gas flowing
through economizer 40, prospectively resulting in an increase in the overall
efficiency of the
refrigeration cycle within chiller 20.
[0058] As illustrated in FIGS. 2 and 6, the coaxial economizer comprises
an inner
housing 184 and an outer housing 186 having a common longitudinal axis. The
outer
housing 186 has an inlet for receiving a fluid from a stage of a multistage
compressor and an
outlet for conveying a fluid to a stage of a multistage compressor.
[0059] The economizer 40 preferably has two chambers: a flow chamber
forming a
fluid flow path about the inner housing and a economizer flash chamber 158 for
flashing fluid
in a liquid state to a gas state. In one embodiment, the economizer 40 has two
chambers
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isolated by two spiraling baffles 154. The number of baffles 154 may vary. The
baffles 154
isolate an economizer flash chamber 158 and a superheat chamber 160.
[0060] The economizer flash chamber 158 contains two phases of fluid, a
gas and a
liquid. The condenser 44 supplies liquid to the economizer flash chamber 158.
[0061] The spiraling baffles 154 depicted in FIG. 6 form a flow passage
156 between
said flash chamber 158 and the flow chamber 160 for conveying flashed gas from
the flash
chamber 158 to the flow chamber 160. The preferred arrangement enables the
flashed gas
conveyed from the flash chamber 158 and the fluid received from the inlet of
the outer
housing 186 to mix along the fluid flow path toward the outlet of the outer
housing 186. In
one embodiment, the spiraling baffles 154 depicted in FIG. 6 form a flow
passage 156
defined by two injection slots. The flow passage 156 can take other forms,
such as a plurality
of perforations in the baffle 154.
[0062] During operation, gas in the economizer flash chamber 158 is drawn
out
through the injection slots 156 into the superheat chamber 160. The spiraling
baffles 154 are
oriented so that the fluid exits through the two injection slots 156 of the
spiraling baffles 154.
The fluid exits in approximately the same tangential directions as the flow
discharged from
the non-final stage compressor 26. The face areas of the flow passage 156 are
sized to
produce approximately matching velocities and flow rates in the flow passage
156 relative to
the adjacent local mixing superheat chamber 160 (suction pipe side). This
requires a
different injection face area of the flow passage 156 based on the location of
the tangential
discharge conic flow, where a smaller gap results closest to the shortest path
length distance,
and a larger gap at the furthest path length distance. Intermediate superheat
chambers 160
and flash chambers may be provided, for example, when more than two stages of
compression are used.
[0063] The economizer flash chamber 158 introduces approximately 10
percent
(which can be more or less) of the total fluid flow through the chiller 20.
The economizer
flash chamber 158 introduces lower temperature economizer flash gas with
superheated gas
from the discharge conic of the non-final stage compressor 26. The coaxial
economizer 42
arrangement generously mixes the inherent local swirl coming out of the
economizer flash
chamber 158 and the global swirl introduced by the tangential discharge of the
non-final
stage compressor 26 ¨ discharge which, in one embodiment, is typically over
the top of the
outside diameter condenser 44 and the inside diameter of coaxially arranged
economizer 42.
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[0064] The liquid in chamber 162 is delivered to the evaporator 22. This
liquid in the
bottom portion of the economizer flash chamber 158 is sealed from the
superheat chamber
160. Sealing of liquid chamber 162 can be sealed by welding the baffle 154 to
the outer
housing of the coaxially arranged economizer 42. Leakage is minimized between
other
mating surfaces to less than about 5 percent.
[0065] In addition to combining multiple functions into one integrated
system, the
coaxial economizer 40 produces a compact chiller 20 arrangement. The
arrangement is also
advantageous because the flashed fluid from the economizer flash chamber 158
better mixes
with the flow from the non-final stage compressor 26 than existing economizer
systems,
where there is a tendency for the flashed economizer gas not to mix prior to
entering a final
stage compressor 28. In addition, the coaxial economizer 40 dissipates local
conic discharge
swirl as the mixed out superheated gas proceeds circumferentially to the final
stage
compressor 28 to the tangential final stage suction inlet 52. Although some
global swirl does
exist at the entrance to the final stage suction pipe 52, the coaxial
economizer 40 reduces the
fluid swirl by about 80 percent compared to the non-final stage compressor 26
conic
discharge swirl velocity. Remaining global swirl can be optionally reduced by
adding a swirl
reducer or deswirler 146 in the final stage suction pipe 52.
[0066] Turning to FIG. 6, a vortex fence 164 may be added to control
strong localized
corner vortices in a quadrant of the conformal draft pipe 142. The location of
the vortex
fence 164 is on the opposite side on the most tangential pick up point of the
coaxially
arranged economizer 42 and the conformal draft pipe 142. The vortex fence 164
is preferably
formed by a sheet metal skirt projected from the inner diameter of the
conformal draft pipe
142 (no more than a half pipe or 180 degrees is required) and bounds a surface
between the
outside diameter of the condenser 44 and inner diameter of the coaxially
arranged
economizer 42. The vortex fence 164 eliminates or minimizes corner vortex
development in
the region of the entrance of the draft pipe 142. The use of a vortex fence
164 may not be
required where a spiral draft pipe 142 wraps around a greater angular distance
before feeding
the inlet flow conditioning assembly 54.
[0067] From the coaxial economizer 40 of this embodiment, the refrigerant
vapor is
drawn by final stage impeller 58 of the final stage compressor 28 and is
delivered into a
conformal draft pipe 142. Referring to FIG. 5, the conformal draft pipe 142
has a total pipe
wrap angle of about 180 degrees, which is depicted as starting from where the
draft pipe 142
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changes from constant area to where it has zero area. The draft pipe exit 144
of the draft
pipe 142 has an outside diameter surface that lies in the same plane as the
inner diameter of
the condenser 44 of the coaxially arranged economizer 42. Conformal draft pipe
142
achieves improved fluid flow distribution, distortion control and swirl
control entering a
later stage of compression.
[0068] Conformal draft pipe 142 can have multiple legs. Use of multiple
legs may
be less costly to produce than a conformal draft pipe 142 as depicted in FIG.
5. Use of such
a configuration has a total pipe wrap angle that is less than 90 degrees,
which starts from
about where projected pipe changes from constant area to a much reduced area.
A draft
pipe 142 with multiple legs achieves about 80 percent of the idealized pipe
results for
distribution, distortion and swirl control.
[0069] Referring still to FIG. 6, fluid is delivered from the draft pipe
142 to a final
stage suction pipe 52. The final stage suction pipe 52 is similarly, if not
identically,
configured to the inlet suction pipe 50. As discussed, the suction pipe 50, 52
can be a three-
piece elbow. For example, the illustrated final suction pipe 52 has a first
leg 52A, section
leg 52B, and a third leg 52C.
[0070] Optionally, a swirl reducer or deswirler 146 may be positioned
within the
final stage suction pipe 52. Details of the structure, function and operation
of a preferred
swirl reducer 146 are disclosed in U.S. patent nos. 8,037,713 and 7,856,834
and U.S. patent
application publication no. 2009/0208331, commonly assigned to the assignee of
the
present invention. A brief discussion of a preferred swirl reducer 146
follows; however,
other swirl reducers may be used with embodiments of the present invention.
[0071] The swirl reducer 146 may be positioned in the first leg 52A,
second leg
52B, or third leg 52C. Referring to FIG. 6, an embodiment of the swirl reducer
146 has a
flow conduit 148 and radial blades 150 connected to the flow conduit 148 and
the suction
pipe 50, 52. The number of flow conduits 148 and radial blades 150 varies
depending on
design flow conditions. The flow conduit 148 and radial blade 150, cambered or
uncambered, form a plurality of flow chambers 152. The swirl reducer 146 is
positioned
such that the flow chambers 152 have a center coincident with the suction pipe
50, 52. The
swirl reducer 146 swirling upstream flow to substantially axial flow
downstream of the
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swirl reducer 146. The flow conduits 148 preferably have two concentric flow
conduits 148
and are selected to achieve equal areas and minimize blockage.
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[0072] The number of chambers 152 is set by the amount of swirl control
desired.
More chambers and more blades produce better deswirl control at the expense of
higher
blockage. In one embodiment, there are four radial blades 150 that are sized
and shaped to
turn the tangential velocity component to axial without separation and provide
minimum
blockage.
[0073] The location of the swirl reducer 146 may be located elsewhere in
the suction
pipe 52 depending on the design flow conditions. As indicated above, the swirl
reducer 146
may be placed in the non-final stage suction pipe 50 or final stage suction
pipe 52, in both
said pipes, or may not be used at all.
[0074] Also, the outside wall of the swirl reducer 146 can coincide with
the outside
wall of the suction pipe 52 and be attached. Alternatively, the one or more
flow conduits 148
and one or more radial blades 150 can be attached to an outside wall and
inserted as a
complete unit into suction pipe 50, 52.
[0075] As illustrated in FIG. 6, a portion of radial blade 150 extends
upstream beyond
the flow conduit 148. The total chord length of the radial blade 150 is set in
one embodiment
to approximately one-half of the diameter of the suction pipe 50, 52. The
radial blade 150
has a camber roll. The camber roll of the radial blade 150 rolls into the
first about forty
percent of the radial blade 150. The camber roll can vary. The camber line
radius of
curvature of the radial blade 150 is set to match flow incidence. One may
increase incidence
tolerance by rolling a leading edge circle across the span of the radial blade
150.
[0076] The radial uncambered portion of the radial blade 150 (no
geometric turning)
is trapped by the concentric flow conduits 148 at about sixty percent of the
chord length of
the radial blade 150. The refrigerant exits the swirl reducer 146 positioned
in the final stage
suction pipe 52 and is further drawn downstream by the final stage compressor
28. The fluid
is compressed by the final stage compressor 28 (similar to the compression by
the non-final
stage compressor 26) and discharged through the external volute 62 out of a
final stage
compressor outlet 34 into condenser 44. Referring to FIG. 2, the conic
discharge from the
final stage compressor 28 enters into the condenser approximately tangentially
to the
condenser tube bundles 46.
[0077] Turning now to the condenser 44 illustrated in FIGS. 1-3 and 6,
condenser 44
can be of the shell and tube type, and is typically cooled by a liquid. The
liquid, which is
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typically city water, passes to and from a cooling tower and exits the
condenser 44 after
having been heated in a heat exchange relationship with the hot, compressed
system
refrigerant, which was directed out of the compressor assembly 24 into the
condenser 44 in a
gaseous state. The condenser 44 may be one or more separate condenser units.
Preferably,
condenser 44 may be a part of the coaxial economizer 40.
[0078] The heat extracted from the refrigerant is either directly
exhausted to the
atmosphere by means of an air cooled condenser, or indirectly exhausted to the
atmosphere
by heat exchange with another water loop and a cooling tower. The pressurized
liquid
refrigerant passes from the condenser 44 through an expansion device such as
an orifice (not
shown) to reduce the pressure of the refrigerant liquid.
[0079] The heat exchange process occurring within condenser 44 causes the
relatively
hot, compressed refrigerant gas delivered there to condense and pool as a
relatively much
cooler liquid in the bottom of the condenser 44. The condensed refrigerant is
then directed
out of condenser 44, through discharge piping, to a metering device (not
shown) which, in a
preferred embodiment, is a fixed orifice. That refrigerant, in its passage
through metering
device, is reduced in pressure and is still further cooled by the process of
expansion and is
next delivered, primarily in liquid form, through piping back into evaporator
22 or
economizer 42, for example.
[0080] Metering devices, such as orifice systems, can be implemented in
ways well
known in the art. Such metering devices can maintain the correct pressure
differentials
between the condenser 42, economizer 42 and evaporator 22 of the entire range
of loading.
[0081] In addition, operation of the compressors, and the chiller system
generally, is
controlled by, for example, a microcomputer control panel 182 in connection
with sensors
located within the chiller system that allows for the reliable operation of
the chiller, including
display of chiller operating conditions. Other controls may be linked to the
microcomputer
control panel, such as: compressor controls; system supervisory controls that
can be coupled
with other controls to improve efficiency; soft motor starter controls;
controls for regulating
guide vanes 100 and/or controls to avoid system fluid surge; control circuitry
for the motor or
variable speed drive; and other sensors/controls are contemplated as should be
understood. It
should be apparent that software may be provided in connection with operation
of the
variable speed drive and other components of the chiller system 20, for
example.
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=
[0082] It will be readily apparent to one of ordinary skill in the art that
the centrifugal
chiller disclosed can be readily implemented in other contexts at varying
scales. Use of
various motor types, drive mechanisms, and configurations with embodiments of
this
invention should be readily apparent to those of ordinary skill in the art.
For example,
embodiments of multi-stage compressor 24 can be of the direct drive or gear
drive type
typically employing an induction motor.
[0083] Chiller systems can also be connected and operated in series or in
parallel (not
shown). For example, four chillers could be connected to operate at twenty
five percent
capacity depending on building load and other typical operational parameters.
[0084] While particular features, embodiments, and applications of the
present
invention have been shown and described, including the best mode, other
features,
embodiments or applications may be possible, as will be understood by one of
ordinary
skill in the art. The scope of the invention, therefore, is defined by the
claims.
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